Synchronous drive apparatus and methods

ABSTRACT

A synchronous drive apparatus and method, wherein the apparatus comprises a plurality of rotors comprising at least a first and a second rotor. The first rotor has a plurality of teeth for engaging the engaging sections of an elongate drive structure, and the second rotor has a plurality of teeth for engaging the engaging section of the elongate drive structure. A rotary load assembly is coupled to the second rotor. The elongate drive structure engages about the first and second rotors. The first rotor is arranged to drive the elongate drive structure and the second rotor is arranged to be driven by the elongate drive structure. One of the rotors has a non-circular profile having at least two protruding portions alternating with receding portions. The rotary load assembly is such as to present a periodic fluctuating load torque when driven in rotation, in which the angular positions of the protruding and receding portions of the non-circular profile relative to the angular position of the second rotor, and the magnitude of the eccentricity of the non-circular profile, are such that the non-circular profile applies to the second rotor an opposing fluctuating corrective torque which reduces or substantially cancels the fluctuating load torque of the rotary load assembly.

This application is a continuation application of Ser. No. 11/101,597,filed Apr. 8, 2005, which claims benefit of and is a divisional ofapplication Ser. No. 10/294,933, filed Nov. 15, 2002, which claimspriority to U.S. Provisional Application Nos. 60/333,118, filed Nov. 27,2001 and 60/369,558, filed Apr. 4, 2002, the entirety of each of thethese applications is hereby incorporated into the present applicationby reference thereto, respectively.

FIELD OF INVENTION

The present invention relates to a synchronous drive apparatus, a methodof operating a synchronous drive apparatus and a method of constructinga synchronous drive apparatus. The invention relates to the eliminationor reduction of mechanical vibrations, in particular but not exclusivelyin internal combustion engines.

BACKGROUND OF INVENTION

Synchronous drive systems, such as timing belt-based systems, are widelyused in motor vehicles, as well as in industrial applications. In motorvehicles, for example, timing belts or chains are used to drive thecamshafts that open and close the engine intake and exhaust valves. Alsoother devices such as water pumps, fuel pumps etc. can be driven by thesame belt or chain.

Internal combustion engines produce many types of mechanical vibrationsduring their operation, and these vibrations are usually transmittedthrough the timing belt or chain in the synchronous drive system. Aparticularly intense source of mechanical vibrations is given by theintake and exhaust valves and the camshafts that open and close thoseintake and exhaust valves. Opening and closing the intake and exhaustvalves leads to a type of vibration known as torsional vibration. Whenthe frequency of these vibrations is close to natural frequency of thedrive, system resonance occurs. In resonance the torsional vibrationsand the span tension fluctuations are at their maximum.

As flexible mechanical structures, timing belts and chains areparticularly susceptible to the deleterious effects of mechanicalvibrations. Mechanical vibrations transmitted through the timing belt orchain cause fluctuations in belt or chain tension, which can lead toincreased wear and reduced belt or chain life. Vibrations may also causetiming errors, and result in undesirable amounts of noise.

Conventional techniques to attenuate the vibrations include increasingthe tension on the belt or chain and installing camshaft dampers.Camshaft dampers connect a source of inertia to a camshaft sprocket by avibration-absorbing rubber or silicone. However, increasing the belt orchain tension increases the noise level and reduces the useful life ofthe belt or chain. Installing camshaft dampers is also an undesirablesolution, because of their cost and/or because of lack of space.

In DE-A-195 20 508 (Audi AG), there is disclosed a wrapped belt drivefor an internal combustion engine, the timing belt being wrapped aroundtwo driven pulleys coupled to the camshaft of the engine, and one drivepulley coupled to the crankshaft of the engine. The objective of theinvention is to counter the torsional vibrations which are found in suchbelt drives. It is proposed to provide an additional torsional vibrationthrough which the critical resonance can be moved to a range where itcan either be tolerated, or does not arise. It is proposed in thecitation to produce torsional vibrations by an “out of round” pulley,which is shown as consisting of one of the camshaft pulleys. The out ofround pulley which is shown has four protruding portions and fourreceding portions arranged regularly around the pulley. It is said thatthe variations in the pulley profile introduce torsionals to the timingbelt at the incoming or outgoing spans of the driven pulleys, which aresuperimposed on the dynamics of the combustion engine, and thus shift oreliminate the critical resonance range. A figure is shown which is saidto show a graph of torsional vibrations of the timing drive in degreescamshaft over the RPM of the crankshaft. The total amplitude is shown,and also the dominant vibration of the second order and the lessrelevant vibrations of the fourth order are shown. A single example of amagnitude of eccentricity of an out of round pulley is given, but noteaching is given as to how to select the magnitude of the eccentricity,and the angular alignment of the out of round rotor relative to theother rotors, for any given conditions of type of engine, type of drivebelt, and type of load. As has been mentioned, the objective of theinvention in the citation is to counter the torsional vibrations in thebelt drive, and not to deal with the source of the vibrations.

In Japanese Utility Model JP 62-192077 (Patent Bulletin No. HEI 1-95538)of 1987 (Hatano et al/Mitsubishi), there is disclosed a tensionequalising driving device which transmits the rotation of a drive pulleyto a driven pulley by a belt drive such as a timing belt in an internalcombustion engine. There is shown a timing belt arrangement in which atoothed pulley of the drive shaft of a camshaft is driven by an ovaltiming belt driving sprocket connected to the drive shaft of an internalcombustion engine. The teaching of the citation is that the drive pulleyis made oval in shape so as to give the drive belt a tension fluctuationwith a phase opposite to that of the tension fluctuation in the beltproduced by the rotation of the internal combustion engine. It is saidthat the drive pulley is installed in such a way that it gives the drivebelt a tension fluctuation with a phase opposite to that of the tensionfluctuation of the belt already present. The oval drive sprocket is saidto be a tension equalising device, and is provided to equalise thetension in the drive belt. A figure is shown of a graph illustrating thetension caused by the valve train torque and the tension caused by thetension equalising device (the oval drive sprocket), the two tensionsbeing shown of the same magnitude and opposite phase. There is nospecific teaching given as to how to determine the magnitude of theeccentricity of the oval drive pulley, nor how to relate the angularposition of the drive pulley to the camshaft pulley which is driven bythe belt. In addition, as discussed in Japanese Application No. HEI9-73581 (Patent Bulletin No. HEI 10-266868) of 1997 (Kubo/Mitsubishi),it was subsequently determined by the Applicant in JP 62-192077 (HEI1-95538) that the use of an oval sprocket as a crank sprocket has anumber of difficulties and problems and is thus not desirable.

SUMMARY OF INVENTION

In accordance with the present invention in a first aspect, there isprovided a synchronous drive apparatus, comprising a continuous-loopelongate drive structure having a plurality of engaging sections. Aplurality of rotors comprising at least a first and a second rotor,wherein the first rotor has a plurality of teeth for engaging theengaging sections of the elongate drive structure, and the second rotorhas a plurality of teeth for engaging the engaging section of theelongate drive structure. A rotary load assembly is coupled to thesecond rotor. The elongate drive structure engages about the first andsecond rotors. The first rotor is arranged to drive the elongate drivestructure and the second rotor is arranged to be driven by the elongatedrive structure. One of the rotors has a non-circular profile having atleast two protruding portions alternating with receding portions. Therotary load assembly is such as to present a periodic fluctuating loadtorque when driven in rotation, in which the angular positions of theprotruding and receding portions of the non-circular profile relative tothe angular position of the second rotor, and the magnitude of theeccentricity of the non-circular profile, are such that the non-circularprofile applies to the second rotor an opposing fluctuating correctivetorque which reduces or substantially cancels the fluctuating loadtorque of the rotary load assembly.

In preferred forms of the apparatus, the non-circular profile is such asto produce the opposing fluctuating corrective torque by periodicelongation and contraction of the spans of the elongate drive structureadjoining the rotor on which the non-circular profile is formed. Theelongate drive structure has a drive span on the tight side of the rotoron which the non-circular profile is formed, the angular position of thenon-circular profile being within +/−15 degrees (preferably within +/−5degrees) of an angular position for which a maximum elongation of thedrive span coincides with a peak value of the fluctuating load torque ofthe rotary load assembly. Most preferably the angular position of thenon-circular profile is that for which a maximum elongation of the drivespan substantially coincides with a peak value of the fluctuating loadtorque of the rotary load assembly.

Also in preferred forms of the apparatus, the magnitude of theeccentricity of the non-circular profile is such that the fluctuatingcorrective torque has an amplitude in the range of 70% to 110%(preferably in the range 90% to 100%) of the amplitude of thefluctuating load torque at a predetermined selected set of operatingconditions of the synchronous drive apparatus. Most preferably, theamplitude of the fluctuating corrective torque is substantially equal tothe amplitude of the fluctuating load torque.

In this specification, unless otherwise indicated, the term amplitude ofa periodically varying item means peak-to-peak amplitude.

Thus, the magnitude of the eccentricity of the non-circular profile isdetermined with reference to the amplitude of the fluctuating loadtorque of the rotary load assembly. In some arrangements the amplitudeof the fluctuating load torque may be substantially constant, and inother arrangements the amplitude of the fluctuating load torque mayvary. Where the amplitude of the fluctuating load torque is constant,the magnitude of the eccentricity is determined with reference to thatsubstantially constant amplitude of fluctuating load torque. Where theamplitude of the fluctuating load torque varies, the value thereof whichis used to determine the magnitude of the eccentricity will be selectedaccording to the operating conditions in which it is desired toeliminate or reduce the unwanted vibrations. For example where thefluctuating load torque of the rotary load assembly varies, theeccentricity may be determined with reference to the amplitude of thefluctuating load torque when determined at conditions such that it is amaximum, or for example when determined at the natural resonancefrequency of the apparatus. For example in a diesel internal combustionengine, the most troublesome region for vibration may be at the maximumfuel delivery by the fuel pump. In these conditions, the eccentricity isdetermined with reference to the amplitude of the fluctuating loadtorque when determined at these conditions. Similarly in a petrol orgasoline internal combustion engine, the most troublesome region may beat the region of natural resonance of the timing drive, and in such acase the eccentricity is determined with reference to such conditions.

It is to be appreciated that the invention finds application in manyforms of synchronous drive apparatus other than in internal combustionengines. Also, the non-circular profile may be provided in manydifferent locations within the drive apparatus. For example anon-circular profile may be provided on the first rotor (which drivesthe elongate drive structure), and/or on the second rotor (which isdriven by the elongate drive structure), and/or may be provided on athird rotor, for example an idler rotor urged into contact with thecontinuous loop elongate drive structure.

However, the invention finds particular use when installed in aninternal combustion engine and the first rotor comprises a crankshaftsprocket. In some arrangements the internal combustion engine is adiesel engine, and the rotary load assembly comprises a rotary fuelpump. As has been mentioned in such arrangements, it may be arrangedthat the magnitude of the eccentricity of the non-circular profile issuch that the fluctuating corrective torque has an amplitudesubstantially equal to the amplitude of the fluctuating load torque whendetermined at conditions of maximum delivery of the fuel pump. In otherarrangements, the internal combustion engine may be a petrol or gasolineengine and the rotary load assembly may be a camshaft assembly.

In determining the angular position of the non-circular profile,consideration may be given to various reference parameters of theprofile and the rotor on which it is formed. In some arrangements thenon-circular profile has at least two reference radii, each referenceradius passing from the centre of the rotor on which the non-circularprofile is formed and through the centre of a protruding portion of thenon-circular profile, and the angular position of the non-circularprofile is related to a reference direction of the rotor on which thenon-circular profile is formed, the reference direction being thedirection of the hub load force produced by engagement of the elongatedrive structure with that rotor. The angular position of thenon-circular profile is such that, when the fluctuating load torque ofthe rotary load assembly is at a maximum, the annular position of areference radius is preferably within a range of 90° to 180° from thereference direction taken in the direction of rotation of the rotor onwhich the non-circular profile is formed. Preferably, the rangecomprises a range of 130° to 140°. Most preferably, the angular positionof the reference radius is substantially at 135° from the referencedirection taken in the direction of rotation of the rotor on which thenon-circular profile is formed.

It will be appreciated that many different forms of non-circular profilemay be provided, for example a generally oval profile, or a profilehaving three or four protruding portions arranged regularly around therotor. The choice of profile will depend upon other components of thesynchronous drive apparatus. Examples which may be provided include thefollowing, namely: the internal combustion engine is a 4-cylinder inlinecombustion engine and the crankshaft sprocket has an oval contouredprofile; the internal combustion engine is a 4-cylinder inlinecombustion engine and the camshaft sprocket has a generally rectangularcontoured profile; the internal combustion engine is a 4-cylinder inlinecombustion engine, and the camshaft sprocket has a generally rectangularcontoured profile and the crankshaft sprocket has an oval contouredprofile; the internal combustion engine is a 3-cylinder inlinecombustion engine and the camshaft sprocket has a generally triangularcontoured profile; the internal combustion engine is a 6-cylinder inlinecombustion engine and the crankshaft sprocket has a generally triangularcontoured profile; the internal combustion engine is a 6-cylinder V6combustion engine and the camshaft sprocket has a generally triangularcontoured profile; the internal combustion engine is an 8-cylinder V8combustion engine and the camshaft sprocket has a generally rectangularcontoured profile; or the internal combustion engine is a 2-cylindercombustion engine and the camshaft sprocket has an oval contouredprofile.

In most embodiments of the invention as set out above, the protrudingportions and receding portions will be generally of the same magnitude,giving a regular non-circular profile. However depending upon thecircumstances of the torsional vibrations to be removed, a non-regularprofile may be provided. Furthermore, the protruding portions referredto above may constitute major protruding portions and the recedingportions constitute major receding portions, and the non-circularprofile may include additional minor protruding portions of lesserextent than the major protruding portions. These minor protrudingportions may be adapted to produce additional, minor, fluctuatingcorrective torque patterns in the torque applied to the second rotor,for the purpose of reducing or substantially cancelling subsidiary orderfluctuating load torque presented by the rotary load assembly, inparticular for example in order to reduce or substantially cancel fourthorder fluctuating load torques presented by the rotary load assembly.

It is to be appreciated that where features of the invention are set outherein with regard to apparatus according to the invention, suchfeatures may also be provided with regard to a method according to theinvention (namely a method of operating a synchronous drive apparatus,or a method of constructing a synchronous drive apparatus), and viceversa.

In particular, there is provided in accordance with another aspect ofthe invention a method of operating a synchronous drive apparatus whichcomprises a continuous-loop elongate drive structure having a pluralityof engaging sections. A plurality of rotors comprises at least a firstand a second rotor. The first rotor has a plurality of teeth engagingthe engaging sections of the elongate drive structure, and the secondrotor has a plurality of teeth engaging the engaging section of theelongate drive structure. A rotary load assembly is coupled to thesecond rotor. One of the rotors has a non-circular profile having atleast two protruding portions alternating with receding portions. Therotary load assembly presents a periodic fluctuating load torque whendriven in rotation.

The method comprises the steps of engaging the elongate drive structureabout the first and second rotors, driving the elongate drive structureby the first rotor, and driving the second rotor by the elongate drivestructure, and applying to the second rotor by means of the non-circularprofile an opposing fluctuating corrective torque which reduces orsubstantially cancels the fluctuating load torque of the rotary loadassembly.

In accordance with yet another aspect of the invention, there may beprovided a method of constructing a synchronous drive apparatus,comprising:

-   -   (i) assembling components comprising a continuous-loop elongate        drive structure having a plurality of engaging sections, a        plurality of rotors comprising at least a first and a second        rotor, the first rotor having a plurality of teeth for engaging        the engaging sections of the elongate drive structure, and the        second rotor having a plurality of teeth for engaging the        engaging section of the elongate drive structure, and a rotary        load assembly coupled to the second rotor; and    -   (ii) engaging the elongate drive structure about the first and        second rotors, the first rotor being arranged to drive the        elongate drive structure and the second rotor being arranged to        be driven by the elongate drive structure, and one of the rotors        having a non-circular profile having at least two protruding        portions alternating with receding portions, the rotary load        assembly being such as to present a periodic fluctuating load        torque when driven in rotation; and    -   (iii) determining the angular positions of the protruding and        receding portions of the non-circular profile relative to the        angular position of the second rotor, and the magnitude of the        eccentricity of the non-circular profile, to be such that the        non-circular profile applies to the second rotor an opposing        fluctuating corrective torque which reduces or substantially        cancels the fluctuating load torque of the rotary load assembly.

In a preferred form of the method of constructing the synchronous driveapparatus, the method includes:

-   -   (i) arranging the non-circular profile to produce the opposing        fluctuating corrective torque by periodic elongation and        contraction of the spans of the elongate drive structure        adjoining the rotor on which the non-circular profile is formed,        the elongate drive structure having a drive span between the        rotor on which the non-circular profile is formed and the second        rotor, the drive span being positioned on the tight side of the        rotor on which the non-circular profile is formed; and    -   (ii) determining the angular positions of the protruding and        receding portions of the non-circular profile by arranging the        angular position of the non-circular profile to be within +/−15        degrees of an angular position for which a maximum elongation of        the drive span coincides with a peak value of the fluctuating        load torque of the rotary load assembly.

Also in a preferred form of the invention the method of constructing asynchronous drive apparatus includes determining the magnitude of theeccentricity of the non-circular profile is determined by the followingsteps:

-   -   (i) measuring the amplitude of the fluctuating load torque of        the rotary load assembly at a predetermined selected set of        operating conditions of the synchronous drive apparatus;    -   (ii) calculating the required amplitude of periodic elongation        and contraction of the drive span by the following formula:        $L = \frac{T}{rk}$    -   L=the amplitude of the periodic elongation and contraction of        the said drive span;    -   T=the amplitude of the fluctuating load torque of the rotary        load assembly at a predetermined selected set of operating        conditions of the synchronous drive apparatus;    -   r=the radius of the second rotor:    -   k=the stiffness coefficient of the elongate drive structure        defined as $\begin{matrix}        {k = {dF}} \\        {d\quad L}        \end{matrix}$        -   where dF is the force required to produce an increase of            length dL in the length of the structure.    -   (iii) producing and recording data to relate empirically a        series of values of (a) the divergence from circular of the        protruding and receding portions of the non-circular profile        and (b) the resulting amplitude of the periodic elongation and        contraction of the drive span; and    -   (iv) selecting from the data the corresponding eccentricity to        give the required amplitude of the periodic elongation and        contraction of the drive span.

The present invention arises from an understanding that the best way toeliminate or reduce torsional vibrations in a synchronous drive systemis to arrange a non-circular profile on one of the rotors which is suchas to cancel or reduce the fluctuating load torque in the load assembly,rather than trying to cancel or reduce the varying tension in thecontinuous loop drive structure, as was attempted in the prior art.Indeed it is found essential to provide a varying tension in theelongate drive structure, in order to cancel or reduce the fluctuatingload torque in the load assembly. The present invention allows thecancellation, or reduction, of the source of the torsional excitation,rather than endeavouring to deal with the effects of torsionals bycancelling variations in tension in the elongate drive structure.

Thus although it has been known to provide a non-circular profile on oneof the rotors in a synchronous drive assembly, the methods chosen todetermine the magnitude of the eccentricity, and the timing of theprotruding and receding portions of the non-circular profile, have notbeen such as to produce the required result. By way of example, in atypical internal combustion engine, if the eccentricity is chosen suchas to try to equalise the tension in a drive belt, the eccentricity willtypically be considerably too great to cancel the torsional vibrationsin the load assembly. In a typical international combustion engine,there will be a resonant frequency at, say, 2000 to 2500 rpm. If theeccentricity of the non-circular profile is chosen to attempt to cancelany tension variation in the drive belt in the region of resonance, thentypically the eccentricity will be set at much more tension than isrequired to cancel the vibrations. The result will be excessive wear inthe drive belt and the various sprockets, and also the system will notbe successful in reducing vibration.

Considering another manner in which the prior art arrangements weredeficient, it is important to arrange the timing (translated intoangular position) of the non-circular profile, to be correctly relatedto the timing (translated into angular positioning) of the fluctuationsin load torque in the load assembly. Conveniently the relative timing ofthe non-circular profile and the fluctuating load torque of the rotaryload assembly is determined in relation to a periodic elongation andcontraction of a drive span of the elongate drive structure between thefirst and second rotors on the tight side of the first rotor. The mostpreferable arrangement in accordance with the invention is that theangular position of the non-circular profile is that for which a maximumelongation of the drive span of the elongate drive structuresubstantially coincides with a peak value of the fluctuating load torqueof the rotary load assembly. However, the invention can providesubstantial reduction in vibration if the timing is set within a rangeof plus/minus 15° of the preferred angular position. A particularlypreferred range is plus/minus 5° of the preferred angular position.

In contrast, in the prior art it has been attempted to set theeccentricity of the non-circular profile with reference to the tensionin the elongate drive structure. However in a typical internalcombustion engine the peak tension in the drive belt varies in itstiming according to the region of the rpm range which is examined.Typically the peak tension in the drive belt occurs at one timing stagefor the resonant frequency of the engine, and occurs at an earliertiming in the cycle for the rev range below resonance, and occurs at alater part of the timing cycle for the region of the rev range above theresonant condition. Thus, depending upon which conditions are selectedin the prior art in order to attempt to equalise the tension in thedrive belt, the timing of the eccentricity of the non-circular profilemay be ahead of, or may lag behind, the preferred position forcancelling the fluctuating load torque in the load assembly.

Thus to summarise, the present invention provides for the correctselection of the eccentricity and the timing of the non-circularprofile, to be that which most advantageously cancels or reduces thefluctuating load torque in the load assembly.

DESCRIPTION OF THE DRAWINGS

Embodiments of the invention will now be described by way of examplewith reference to the accompanying drawings in which:

FIG. 1 is a schematic illustration of a synchronous drive apparatus fora motor vehicle internal combustion engine, embodying the invention;

FIG. 2 is an enlarged view of the crankshaft sprocket shown in FIG. 1;

FIG. 3 is a schematic illustration of the synchronous drive apparatus ofan internal combustion engine in DOHC engine configuration;

FIG. 4 a shows a graph of a fluctuating load torque at the camshaft ofan SOHC internal combustion engine and a fluctuating corrective torquegenerated by an oval crankshaft sprocket illustrated in FIGS. 1 and 2,all graphs being taken over one crankshaft revolution;

FIG. 4 b shows a graph of a fluctuating load torque which arises fromthe intake cam of an DOHC internal combustion engine, a fluctuating loadtorque which arises from the exhaust cam, and a fluctuating correctivetorque generated by an oval crankshaft sprocket in the engineillustrated in FIG. 3, all graphs being taken over one crankshaftrevolution;

FIGS. 5 a to 5 d show different combinations of crankshaft and camshaftsprockets embodying the invention in 4-cylinder and 3-cylinder engines;

FIGS. 6 a to 6 d show different combinations of crankshaft and camshaftsprockets embodying the invention in 6-cylinder, 8-cylinder and2-cylinder engines;

FIG. 7 a is a graph illustrating the magnitude of torsional vibrationsin an internal combustion engine at different engine speeds, thevertical axis indicating the amplitude of torsional vibrations indegrees of movement of the camshaft, and the horizontal axis indicatingengine speed in rpm, the graph indicating the situation in a knownengine, having a round crankshaft sprocket;

FIG. 7 b is a graph illustrating the magnitude of torsional vibrationsin an internal combustion engine at different engine speeds, thevertical axis indicating the amplitude of torsional vibrations indegrees of movement of the camshaft, and the horizontal axis indicatingengine speed in rpm, the graph indicating the situation for asynchronous drive apparatus embodying the invention, utilising an ovalcrankshaft sprocket;

FIG. 8 a is a graph illustrating the magnitude of tensions in aninternal combustion engine at different engine speeds, the vertical axisindicating the amplitude of the belt tension, and the horizontal axisindicating engine speed in rpm, the graph indicating the situation in aknown engine, having a round crankshaft sprocket;

FIG. 8 b is a graph illustrating the magnitude of tensions in aninternal combustion engine at different engine speeds, the vertical axisindicating the amplitude of the belt tension, and the horizontal axisindicating engine speed in rpm, the graph indicating the situation for asynchronous drive apparatus embodying the invention, utilising an ovalcrankshaft sprocket;

FIGS. 9 a and 9 b show respectively the fluctuations in tension in thedrive belt over one revolution of the crankshaft at 1500 RPM, for anengine according to the prior art, having a round crankshaft sprocket,FIGS. 9 a and 9 b showing respectively the belt tension variations onthe tight side and the slack slide of the crankshaft sprocketrespectively;

FIGS. 10 a and 10 b show respectively the fluctuations in tension in thedrive belt over one revolution of the crankshaft at 2500 RPM, for anengine according to the prior art, having a round crankshaft sprocket,FIGS. 10 a and 10 b showing respectively the belt tension variations onthe tight side and the slack slide of the crankshaft sprocketrespectively;

FIG. 11 show respectively the fluctuations in tension in the drive beltover one revolution of the crankshaft at 3500 RPM, for an engineaccording to the prior art, having a round crankshaft sprocket, FIGS. 11a and 11 b showing respectively the belt tension variations on the tightside and the slack slide of the crankshaft sprocket respectively;

FIG. 12 is a three-dimensional graph showing the distribution ofcamshaft torsional vibrations in a known internal combustion enginehaving a round crankshaft sprocket, in which the X-axis indicatesvarious harmonic orders of vibration, the Y-axis indicates engine speedin RPM, and the Z-axis indicates the amplitude of the camshaft torsionalvibrations;

FIG. 13 is a three-dimensional graph showing the distribution ofcamshaft torsional vibrations in an engine embodying the invention andhaving an oval crankshaft sprocket, in which the X-axis indicatesvarious harmonic orders of vibration, the Y-axis indicates engine speedin RPM, and the Z-axis indicates the amplitude of the camshaft torsionalvibrations;

FIG. 14 a shows a graph of fluctuating load torque on a rotary loadassembly such as a camshaft;

FIG. 14 b shows how a non-circular profile 19 may be derived to cancelthe torque fluctuations of FIG. 14 a, in an embodiment of the invention;and

FIGS. 15, 16 and 17 show a computer generated virtual representation ofan oval crankshaft profile embodying the invention, the profile beingstepped on by an angular advance of one tooth in FIG. 16 relative toFIG. 15, and in FIG. 17 relative to FIG. 16.

DESCRIPTION OF THE INVENTION

FIG. 1 is a diagrammatic representation of a synchronous drive apparatusfor a motor vehicle internal combustion engine, embodying the invention.The apparatus comprises a continuous loop elongate drive structure 10,first and second rotors 11 and 12, and further rotors 13, 14 and 17. Thecontinuous loop elongate drive structure 10 is provided by aconventional timing belt having teeth 15 together with interveningvalleys which constitute a plurality of engaging sections of thecontinuous loop elongate drive structure. Each rotor 11 and 12 isprovided by a sprocket having a plurality of teeth 16 for engaging thevalleys between the teeth 15 of the timing belt 10. The sprocket 11 iscoupled to the crankshaft (not shown) of an internal combustion engine,and the sprocket 12 is coupled to a rotary load assembly (not shown)which is constituted by a camshaft 26 of the internal combustion engine.The timing belt 10 is engaged about the first and second rotors 11 and12, the first rotor 11 being arranged to drive the belt 10 and thesecond rotor 12 being arranged to be driven by the belt 10. The rotor 14also has teeth 16 and consists of a sprocket for driving other elementsof the internal combustion engine, such as a water pump, and the rotor13 is preferably for a belt tensioner bearing on a non-toothed side ofthe timing belt 10, to tension the belt in known manner. Rotor 17 ispreferably for a fixed idler pulley bearing on the non-toothed side oftiming belt 10.

In a known form of a synchronous drive apparatus, the crankshaftsprocket would have a circular profile. In such a case, the synchronousdrive apparatus is prone to vibrations, known as torsional vibrations,which arise from the opening and closing of the intake and exhaustvalves of the internal combustion engine by the overhead camshaft. Thesource of the excitations is illustrated in FIGS. 4 a and b. FIG. 4 aillustrates the fluctuating load torque 103 applied to the camshaft in aSOHC engine and FIG. 4 b illustrates the same for a DOHC engine. FIG. 4b shows the variation of camshaft torque over a single cycle of theengine, indicating how the intake torque shown by the curve 101 varieswith degrees of rotation of the engine, and how the exhaust torqueprofile 102 varies in the same way.

In accordance with the embodiment of the present invention shown in FIG.1 for a SOHC engine, the crankshaft sprocket 11 has a non-circularprofile (as shown in exaggerated form in FIG. 2) indicated generally byreference numeral 19. The non-circular profile 19 is, in the particularembodiment described, an oval having a major axis 20 and a minor axis21. The profile 19 has two protruding portions 22 and 23 and has tworeceding portions 24 and 25.

The provision of the oval profile 19 on the sprocket 11 as shown in FIG.2, generates a fluctuating corrective torque, which is applied by thebelt 10 to the second rotor 12. This fluctuating corrective torque isshown at 104 in FIG. 4 a. In the preferred situation, the totalfluctuating load torque 103 is opposed by the overall corrective torque104. Preferably the corrective torque 104 is 180° out of phase with theoverall load torque 103, and the peak to peak amplitude of thefluctuating corrective torque 104 is made equal to the peak to peakamplitude of the overall fluctuating load torque 103.

In accordance with the embodiment of the invention using the ovalprofile 19 shown in FIG. 2, the angular positions of the protruding andreceding portions 22 to 24 of the non-circular profile 19 relative tothe angular position of the second rotor 12, and the magnitude of theeccentricity of the non-circular profile 19, are such that thenon-circular profile 19 applies to the second rotor 12 an opposingfluctuating corrective torque 104 which substantially cancels thefluctuating load torque 103 of the rotary load assembly 26.

The determination of the timing and magnitude of the eccentricity of thenon-circular profile 19 will now be described in more detail. In FIG. 1the spans between the various rotors are indicated as 10A between rotor12 and rotor 14, 10B between rotor 14 and rotor 11, 10C between rotor 12and rotor 13, and 10D between rotor 13 and rotor 17 and 10E betweenrotor 17 and rotor 11. The span between the first rotor 11 and thesecond rotor 12, indicated as 10A, 10B, is referred to as the drive spanbetween the two rotors, it being positioned on the tight side of thefirst rotor 11 on which the non-circular profile 19 is formed. The spanbetween the first rotor 11 and second rotor 12 which is indicated as10C, 10D, 10E is referred as the slack side, although of course the beltis under tension on both sides. The torsional vibrations to beeliminated are formed by the fluctuating load torque on the rotary loadassembly (the camshaft 26) and in accordance with the present inventionthis is reduced or substantially cancelled by the application of anopposing fluctuating corrective torque to the camshaft 26 by means ofthe timing belt 10. The opposing fluctuating corrective torque isproduced by the non-circular profile 19 by periodic elongation andcontraction of the spans 10A 10B and 10C 10D 10E, adjoining the rotor 11on which the non-circular profile is formed. In preferred forms of theinvention, the angular position of the non-circular profile 19 is set asclosely as possible to be that for which a maximum elongation of thedrive span 10A 10B substantially coincides with a peak value of thefluctuating load torque of the camshaft 26. It may not always bepossible to arrange this exactly, and advantage is obtained inaccordance with the invention if the angular position of thenon-circular profile is within +/−15 degrees of the preferred angularposition, more preferably within +/−5 degrees.

With regard to the particular case illustrated, and referring to FIGS. 1and 2, the oval profile 19 has two reference radii 20 a and 20 b, whichtogether form the major axis 20 of the oval. Each reference radius 20 a,20 b passes from the centre of the rotor 11 and through the centre ofthe respective protruding portion 22, 23. The angular position of thenon-circular profile 19 is related to a reference direction of the rotor11, the reference direction being the direction of a vector or imaginaryline 27 that bisects the angle or sector of wrap of the continuous loopdrive structure 10 around the rotor 11. This vector that bisects theangle of wrap is in the same direction as the hub load force produced byengagement of the belt 10 with the rotor 11 when the belt drive systemis static. It should be appreciated, however, that the hub load forcedirection changes dynamically during operation of the belt drive system.The timing of the non-circular profile 19 is set to be such that, at thetime when the fluctuating load torque on the second rotor 12 is at amaximum, the angular position of the reference radius 20 a is within arange of 90° to 180° from the reference direction of the angle of wrapbisection 27, taken in the direction of rotation of the rotor 11,preferably within a range of 130° to 140°. Assuming that the assembly ofFIG. 1 is shown at the instant when the fluctuating load torque on thesecond rotor 12 is at a maximum, then the preferred timing of thenon-circular profile 19 is as shown in FIG. 1, namely that the anglebetween the reference radius 20 a and the bisection direction 27 is135°, as indicated by the angle θ.

It is to be appreciated that in this specification, where the term“reference radius” is used for a non-circular profile 19, the referenceparameter measured is the radius of a notional circle passing throughthe associated protruding portion, and is not a radius of the entireprofile, since this entire profile is essentially non-circular. The termreference radius is used merely to indicate the distance between thecentre of the axis of the rotor on which the profile is formed, to themaximum extent of the profile at the relevant protruding portion.

Consideration will now be given to the determination of the magnitude ofthe eccentricity of the non-circular profile 19 in the specificembodiment shown. In summary, the magnitude of the eccentricity of theprofile 19 is preferably set to be such that the fluctuating correctivetorque 104 shown in 4 a has an amplitude substantially equal to, andphase substantially opposite to, the amplitude of the fluctuating loadtorque 103 shown in FIG. 4 a. However advantage is still found inembodiments where the amplitude of the fluctuating corrective torque 104is in the range of 75% to 110% of the amplitude of the fluctuating loadtorque 103, more preferably in the range 90% to 100%. Where thefluctuating load torque 103 has a substantially constant amplitude overthe rev range of the engine, the amplitude of the corrective torque 104is merely made substantially equal to the constant amplitude of thefluctuating load torque.

The practical steps of determining the magnitude of the eccentricity maybe as follows. First the amplitude of the fluctuating load torque 103 ofthe camshaft 26 is measured at the selected set of operating conditions,in this case at the maximum amplitude of the fluctuating load torque.Next there is calculated the required amplitude of period elongation andcontraction of the drive span 10 a, 10 b by the following formula:$L = \frac{T}{rk}$where:

-   -   L=the amplitude of the periodic elongation and contraction of        the drive span which is required;    -   T=the amplitude of the fluctuating load torque of the camshaft        26, which has been measured at maximum amplitude;    -   r=the radius of the second rotor 12: and    -   k=the stiffness coefficient of the belt 10.

The stiffness coefficient k is obtained from the formula k = dF d  Lwhere dF is the force required to produce an increase of length dL inthe of the structure.

By way of example of the calculations above, the amplitude of thefluctuating load torque T may be 10 Nm (zero to peak), and the radius ofthe rotor 12 may be 50 mm. This gives a maximum force F required toprovide the required fluctuating corrective torque of F=200N. In theexample discussed, the required change in span length is obtained bydividing the tension of 200N by the stiffness coefficient k, which forexample for a typical belt may be 400 N/mm. This gives requiredamplitude of elongation and contraction of the timing belt of 0.5 mm(zero to peak).

The next step is to calculate the eccentricity required to provide thislength of elongation and contraction at a timing stage when the majoraxis 20 of the ellipse is set at θ=135° as shown in FIG. 1. Atheoretical calculation of this value is difficult to achieve, so thatthe calculation of the eccentricity is arrived at by the equivalent of a“look-up” table. This is done by producing and recording data to relateempirically a series of values of (i) the divergence from circular ofthe protruding and receding portions of the non-circular profile and(ii) the resulting amplitude of the periodic elongation and contractionof the drive span. The required eccentricity is then selected from thedata to give the required amplitude of the periodic elongation andcontraction of the drive span.

The data bank which is produced, to provide the “look-up” table consistsof a table of values of the amplitude of elongation and contraction ofthe drive span 10A and 10B, for various values of the eccentricity ofthe oval profile 19 along the major axis. Examples of such data aregiven in the following table, Table 1. The reference circle used forcomparison is a circle having a diameter equal to the average of themajor axis length 20 and the minor axis length 21. The eccentricity ofthe oval profile 19 can be determined, in the example shown, byconsidering the divergence of the outline from the reference circle atthe major axis 20. Difference between Amplitude of periodic selectedoval reference elongation and contraction outline and reference circleof drive span 10A, 10B 0.5 mm 0.25 mm 1.0 mm 0.49 mm 1.5 mm 0.74 mm

This table may be derived for example by producing a computer simulationof the oval profile 19, and stepping this through a series of angularadvancements of, say one tooth at a time, for example as shown in FIGS.15, 16 and 17. For each of these steps, the computer simulation isarranged to provide an indication of the elongation or contraction ofthe equivalent drive span 10A, 10B, for a particular length of majoraxis giving the radius 20A. On the computer simulation, the referenceradius 20A is then varied, and a further series of data are produced forthe new radius 20A. The purpose of stepping the profile through thepositions shown at FIGS. 15, 16 and 17, is to determine empirically theposition at which the maximum extension of the corresponding drive span10A, 10B takes place. Having determined that, the appropriate data isextracted, for the maximum length of the span 10A, 10B, which is setagainst the corresponding eccentricity of the reference radius 20A.FIGS. 15, 16 and 17 show how the amplitude of elongation may bedetermined by using virtual prototyping.

FIGS. 5 a to 5 d show different combinations of crankshaft and camshaftsprockets for 4-cylinder and 3-cylinder engines. FIGS. 6 a to 6 d showdifferent combinations of crankshaft and camshaft sprockets for6-cylinder, 8-cylinder and 2-cylinder engines.

FIG. 7 a shows the amplitude of camshaft torsional vibrations in degreesof rotary vibration versus the engine speed in rpm for a roundcrankshaft sprocket. FIG. 7 b shows the amplitude of camshaft torsionalvibrations in degrees of rotary vibration versus the engine speed in rpmfor an oval crankshaft sprocket. FIG. 7 b shows that the torsionals aresignificantly reduced. Only torsionals coming from the crankshaftremain. The resonance has been cancelled.

FIG. 8 a shows the tight side tension fluctuation versus the enginespeed in rpm for a round crankshaft sprocket. FIG. 8 b shows the tightside tension fluctuation versus the engine speed in rpm for an ovalcrankshaft sprocket. FIG. 8 b also shows that resonance has beencancelled. Tension fluctuations are still present in the whole rpmrange, but they need to be there to provide cancelling torque.

FIGS. 9 a and b show the tight side and slack side tension fluctuationsover one revolution of the round crankshaft sprocket at 1500 rpm. FIGS.10 a and b show the tight side and slack side tension fluctuations overone revolution of the round crankshaft sprocket at the system resonance(2500 rpm). FIGS. 11 a and b show the tight side and slack side tensionfluctuations over one revolution of the round crankshaft sprocket at3500 rpm.

FIG. 12 shows the camshaft torsional vibrations for a round crankshaftsprocket presented as a spectral analysis where: x-axis=harmonicsorders; y-axis=engine rpm; and z-axis=amplitude of the camshafttorsional vibrations.

FIG. 13 shows the camshaft torsional vibrations for an oval crankshaftsprocket presented as a spectral analysis where: x-axis=harmonicsorders; y-axis=engine rpm; and z-axis=amplitude of the camshafttorsional vibrations. Only second order torsionals are eliminated by theoval profile. Using a more complex profile, as shown in FIG. 14 willcancel simultaneously second and fourth order torsionals.

FIGS. 14 a and 14 b show, in greatly exaggerated form, how anon-circular profile 19 of one of the rotors in a synchronous driveapparatus embodying the invention can be shaped to accommodate twodifferent orders of torsional fluctuations in the torque of a rotaryload assembly. FIG. 14 consists of two FIGS. 14 a and 14 b. FIG. 14 ashows in curve 110 a second order fluctuating load torque, equivalent tothe second order peak shown in FIG. 12. The curve 111 shows a fourthorder fluctuating load torque equivalent to the fourth order peak shownin FIG. 12. Curve 112 shows the combined fluctuating load torque on therotary load assembly.

In FIG. 14 b there is shown at 19A in greatly exaggerated form agenerally oval profile suitable for use on a crankshaft rotor 11 in FIG.1, having protruding portions 22 and 23. These protruding portionsproduce a corrective fluctuating load torque which can be applied tocancel the second order fluctuating load torque 110 in FIG. 14 a. Asecond profile indicated at 19B is shaped to have four minor protrudingportions which, if it were to be used as a profile of crankshaftsprocket 11, would produce a corrective torque equivalent to the fourthorder fluctuating load torque 111 in FIG. 14 a. In FIG. 14 b, anon-circular profile embodying the invention is indicated at 19C, whichis a combination of the two profiles 19A and 19B. The combined profile19C has two major protruding portions, and two minor protrudingportions. The combined profile 19C produces a fluctuating correctivetorque which can be made to cancel the combined fluctuating torque 112shown in FIG. 14 a.

Thus in FIG. 14, there is shown a modification of the oval rotor inwhich additional minor protruding portions of the profile are provided.The reason for this is to take account of fourth order harmonictorsional vibrations which are illustrated in FIGS. 12 and 13. In FIG.12, there is shown the torsional vibrations which arise from the second,fourth and sixth order harmonics, with a synchronous drive apparatushaving a circular crankshaft sprocket. FIG. 13 shows the torsionalvibrations remaining after use of an oval crankshaft drive sprocket inaccordance with the invention. It will be seen that the fourth orderharmonic torsional vibrations remain. These vibrations can be reduced oreliminated by providing on the non-circular profile of the crankshaftsprocket additional protruding portions. The minor protruding portionsare of lesser extent than the major protruding portions, and arearranged to produce lesser fluctuating corrective torque patterns in thetorque applied to the second rotor, to reduce or substantially cancelthe fourth order fluctuating load torque presented by the rotary loadassembly.

Returning now to a general consideration of the operation of embodimentsof the invention, it is known to provide in a synchronous drive systemfor an internal combustion engine a crankshaft sprocket of oval profile.The present invention provides for the correct selection of theeccentricity and the timing of the non-circular profile, to be thatwhich advantageously cancels or reduces the fluctuating load torque inthe load assembly, rather than endeavouring to equalise the tension inthe drive belt, has as been done in the prior art arrangements.

The invention can be understood by considering Newton's second law, thatthe presence of an unbalanced force will accelerate an object. Forlinear examples this provides:—Acceleration=Force/MassIn rotary motion:Acceleration=Torque/InertiaIn an ordinary internal combustion engine the torque from the valvetrain or diesel fuel pump fluctuates, causing the speed to fluctuate,causing angular displacement to fluctuate (also known as torsionalvibration). By using an ellipsoidal crankshaft sprocket that is pullingthe belt (at appropriate instant) additional torque can be created thathas such amplitude and phase that the combined torque acting on thecamshaft is zero. Absence of torque means absence of acceleration byfirst Newton's law. Absence of acceleration means absence of speedfluctuations, which means that no torsionals are present.

The opening and closing of the intake and exhaust valves is a source oftorque fluctuations. These torque fluctuations cause the camshaft to beinflicted with speed fluctuations, which in turn, causes angularposition fluctuations otherwise know as torsional vibrations. The bestcure for that behaviour is to attack the cause right at the source byintroducing another torque acting on the camshaft i.e. removing torquefluctuations at the camshaft. One way of doing it is to use the ovalsprocket at the crankshaft. The oval sprocket, while rotating, willintroduce fluctuations of span length i.e. will pull and relieve twotimes per one crankshaft revolution. When the tight side is beingpulled, the slack side is relieved and vice versa. Pulling and relievingthe belt means that a new, additional torque is generated at thecamshaft. If this new torque is of appropriate amplitude and phase itcan balance the first torque from the valve train. Absence of torquefluctuations means absence of speed fluctuations and therefore absenceof torsionals.

In embodiments of the invention, when the torsional vibrations in thecamshaft are eliminated the belt tension still varies. Indeed it is thevariation in tension in the belt, which causes the torsional vibrationsin the camshaft to cease. In the prior art, the objective is said to bethe removal of tension variation in the belt, which is not what isneeded to remove torsional vibration in the camshaft. The object is toremove the variation in speed of the driven sprocket, which is caused byvariation in torque load in the driven sprocket. This is done by varyingthe tension in the belt during the cycle of the driven sprocket. At atime of increase of torque load on the driven sprocket, there must be anincrease in tension in the belt. At moment when increase in tension isrequired the effective length of the span must be increased. This isachieved by having the oval positioned so that the long axis is movingfrom a position perpendicular to the hub load, to position along the hubload direction. At the moment when decrease in tension is required theeffective length of the span must be decreased. This is done while themajor axis moves from vertical to horizontal.

1. An improved rotor for use in a synchronous drive apparatus, saiddrive apparatus comprising a continuous-loop elongate drive structurehaving a plurality of engaging sections; a plurality of rotorscomprising at least a first and a second rotor, the first rotor having aplurality of teeth for engaging the engaging sections of the elongatedrive structure, and the second rotor having a plurality of teeth forengaging the engaging section of the elongate drive structure; a rotaryload assembly coupled to the second rotor; the rotary load assemblybeing such as to present a periodic fluctuating load torque when drivenin rotation; the elongate drive structure being engaged about the firstand second rotors, the first rotor being arranged to drive the elongatedrive structure and the second rotor being arranged to be driven by theelongate drive structure, and one of the rotors having a non-circularprofile having at least two protruding portions alternating withreceding portions; wherein the improved non-circular profile rotor hasat least two reference radii, each reference radius passing from thecentre of the non-circular profile rotor and through the centre of aprotruding portion of the non-circular profile, the angular position ofthe non-circular profile being related to a reference direction, thereference direction being the direction of a vector that bisects anangle about which the elongate drive structure is wrapped about therotor having the non-circular profile, the angular position of areference radius is within a range of 90° to 180° from the referencedirection taken in the direction of rotation of the rotor on which thenon-circular profile is formed.
 2. Apparatus according to claim 1, inwhich the angular position of the reference radius being within a rangeof 130° to 140° from the reference direction taken in the direction ofrotation of the rotor on which the non-circular profile is formed. 3.Apparatus according to claim 1, in which the angular position of thereference radius is substantially at 135° from the reference directiontaken in the direction of rotation of the rotor on which thenon-circular profile is formed.